1. Field of the Invention
The present invention relates to an internally meshing planetary gear structure which is ideal for application in reduction gears or speed-up gears, more particularly, which is ideal for application in reduction gears or speed-up gears which have to meet the requirements of being of small size and providing high output.
2. Description of the Prior Art
Conventional internally meshing planetary gear structures comprising a first shaft, an eccentric body mounted on the circumference of the first shaft, a plurality of external-toothed gears which are mounted eccentrically on the first shaft by means of the eccentric body, an internal-toothed gear which meshes internally with the external-toothed gears, and a second shaft coupled to the external-toothed gears through a means which transmits only the rotational component of the external-toothed gears are widely known.
A prior art example of the structure is shown in FIG. 13 and FIG. 14. In the prior art example, by making the first shaft the input shaft, the second shaft the output shaft, and by fixing the internal-toothed gear, the above-described structure is applied to a reduction gear.
Eccentric bodies 3a, 3b are mounted on an input shaft 1 with a fixed phase difference (in this example 180.degree.). Incidentally, the eccentric bodies 3a, 3b are integrated into one body. Two external-toothed gears 5a, 5b are mounted on the eccentric bodies 3a, 3b, respectively, by means of rollers 4. The external-toothed gears 5a, 5b have a plurality of inner roller holes 6, into which inner pins 7 and inner rollers 8 are inserted.
The circumference of the external-toothed gears 5a, 5b is equipped with external teeth 9, which have trochoidal, circular, or the like tooth profiles. The external teeth 9 and an internal-toothed gear 10 mesh internally. The internal-toothed gear 10 is fixed to a casing 12. To be concrete, the structure of the internal teeth of the internal-toothed gear 10 is such that outer pins 11 are loosely fitted into outer pin holes 13 to allow easy rotation.
The inner pins 7, which pass through the external-toothed gears 5a, 5b, are tightly fixed to a flange 14 of an output shaft 2.
When the input shaft 1 makes one revolution, the eccentric bodies 3a, 3b also make one revolution. One revolution of the eccentric bodies 3a, 3b causes the external-toothed gears 5a, 5b to tend to perform swinging and rotating movements around the input shaft 1. However, as the internal-toothed gear 10 restricts the rotation of the external-toothed gears 5a, 5b, the external-toothed gears 5a, 5b perform almost only swinging movements while internally contacting the internal-toothed gear 10.
If it is assumed that, for instance, the number of teeth of the external-toothed gears 5a, 5b is N and the number of teeth of the internal-toothed gear 10 is N+1, then the difference between the numbers of teeth is 1. Consequently, every revolution of the input shaft 1 causes the external-toothed gears 5a, 5b to undergo a shifting (to make a rotation) of one tooth relative to the internal-toothed gear 10 which is fixed to the casing 12. This means that one revolution of input shaft 1 has been decelerated by -1/N revolutions of the external-toothed gear.
The swinging component of the external-toothed gears 5a, 5b is absorbed by gaps between the inner roller holes 6 and the inner pins 7, and only the rotational component of that is transmitted to the output shaft 2 through the inner pins 7.
Therefore, a reduction of reduction ratio -1/N is achieved.
In the prior art example, the internal-toothed gear of the internally meshing planetary gear structure is fixed, the first shaft is the input shaft and the second shaft is the output shaft. However, a reduction gear can also be designed with the second shaft being fixed, the first shaft being the input shaft and the internal-toothed gear being the (big) output shaft. Further, reversing the inputs and outputs also allows to design a step-up gear.
However, the following problems have been associated with the above-described internally meshing planetary gear structure.
As shown in FIG. 15, when considering loads acting on the input shaft 1 and the output shaft 2, it is apparent from the figure that a rotational load W1 from the input shaft 1 to the output shaft 2 acts at the edge position of a bearing 15b. Furthermore, a load W2 from the external-toothed gears (not shown in FIG. 15) to the output shaft 2 acts on the inner pin 7 as illustrated. In addition, a load W3 from the external-toothed gears 5a, 5b to the input shaft 1 acts on the input shaft 1 as shown.
Therefore, the output shaft 2 is subjected to the loads W1, W2 while being in a cantilever state, because the loads W1, W2, which act on the output shaft 2, are located closer to the input shaft 1 than bearings 16a, 16b. Consequently, the resultant moment causes the output shaft 2 to be inclined at an angle .alpha. to its normal axis center 01.
Also, the moment of the load W3, with the load W3 acting on the input shaft 1, causes the input shaft 1 to be inclined by an angle .beta. from its normal axis center 01, in conjunction with the inclination of the output shaft 2.
Consequently, the input shaft 1 and the output shaft 2 rotate with their axis centers being displaced, which causes abnormal wear, noise and vibration of the reduction gear.
Also, as illustrated in FIG. 16, when an external radial load F acts on the input shaft 1, then, in the same manner as described above, the input shaft 1 inclines by an angle .beta.' from its normal axis center 01, and the output shaft 2 inclines by an angle .alpha.' from its normal axis center 01. This inclination also causes abnormal wear, noise and vibration of the reduction gear.
The above-described inclination of the output shaft 2 or the input shaft 1 results from that the input shaft 1 is supported by bearings 15a, 15b on the input shaft 1, with said input shaft 1 receiving the load W3 from the external-toothed gears 5a, 5b through the eccentric bodies 3a, 3b.
On the other hand, in above-described conventional internally meshing planetary gear structures, the radial load is balanced because the transmission torque is equally distributed on the external-toothed gears 5a, 5b. However, since the two external-toothed gears 5a, 5b do not lie in the same plane, a load acting on each external-toothed gear 5a, 5b causes a moment (a couple of forces) to act on the eccentric bodies 3a, 3b (refer to FIG. 13).
Since the moment acting on the eccentric bodies 3a, 3b, corresponds to the product of the load acting on the external-toothed gears 5a, 5b and the distance between the two external-toothed gears 5a, 5b, reducing the distance between the two external-toothed gears 5a, 5b will reduce the moment acting on the eccentric bodies 3a, 3b. However, the external-toothed gears 5a, 5b are supported by the eccentric bodies 3a, 3b and the rollers 4, and the eccentric bodies 3a, 3b and the rollers 4 are required to have an adequate length. Because their load capacity depends upon their length. Therefore, reducing the distance between the external-toothed gears 5a, 5b has its limitations.